Steam turbine and power generating equipment

ABSTRACT

The present invention solves problems of securing high-temperature strength at elevated temperatures and of preventing steam leakage that arise when high-pressure, high-temperature steam is used for driving a steam turbine, and problems of preventing the occurrence of rubbing due to the suppression of the excessive elongation difference and of minimizing steam leakage from shaft seals. A double-wall casing structure is arranged at an area corresponding to stages from a high-pressure first stage ( 7 ) to a predetermined high-pressure stage arranged on an upstream side of a high-pressure final stage ( 8 ); and a single-wall casing structure is arranged at an area corresponding to stages which follows said predetermined high-pressure stage.

TECHNICAL FIELD

The present invention relates to a casing structure for a steam turbineincluded in a thermal power generation system installed in, for example,a combined power plant, and a power generating system using the steamturbine provided with the casing structure.

BACKGROUND OF THE INVENTION

Recently, many combined-cycle power plants provided with a gas turbineand a steam turbine in combination have been constructed. Generally, theimprovement of steam conditions is directly related with the improvementof the efficiency of a power plant provided with a steam turbine.Therefore, the increase of the pressure and temperature of steam fordriving a steam turbine included in a combined-cycle power generatingsystem has been required to improve the efficiency of the powergenerating system and to enhance the output of the power generatingsystem.

As shown in FIG. 8, a casing 110 of a high-pressure stage 5 of aconventional steam turbine for combined-cycle power generation is asingle-wall casing. Usually, the thickness of the wall of thesingle-wall casing must be increased to improve the pressure withstandstrength when inlet steam pressure is raised. In the event that thepressure and temperature of the steam are raised to improve theefficiency of the steam turbine provided with the conventionalsingle-wall casing and to enhance the output of the same, an increasedpressure stress and an increased thermal stress are induced in thecasing owing to increase in the thickness of the wall of the casing. Thecasing is thus damaged by thermal fatigue or high temperature low-cyclefatigue during the operation, and the operation of the turbine affected.

The risk of steam leakage from the horizontal flange of the casing isincreased by increase in thermal deformation of the casing, resulting inthe marked degradation of the reliability of the steam turbine. Steamleakage involves the direct discharge of high-temperature, high-pressuresteam into the atmosphere, which is fatal to the operation of the steamturbine, and increases the risk of fire and injury.

Since an excessively high thermal stress is induced in the casing havinga thick wall at the start of the steam turbine, it must take a long timefor starting up time of the turbine to reduce the level of the thermalstress. However, in a case, such as a combined-cycle power plant whichis required quick start-up, the extension of the starting up time delaysthe start up of the combined-cycle power plant and increases theoperating cost of the power generating system.

When the output of the steam turbine provided with a conventionalsingle-wall casing structure is increased by raising the pressure andtemperature of the main steam, the casing must be made of 12-Cr steel or9-Cr steel, which has strength at high temperatures but expensive,instead of a conventional low alloy steel. The high material cost of thecasing is a principal factor that increases the cost of the steamturbine.

The linear thermal expansion coefficients of the 12-Cr steel and the9-Cr steel are smaller than those of conventional low alloy steels,typically CrMoV steels. Therefore, the thermal expansion of a casingmade of 12-Cr steel or 9-Cr steel is smaller than that of theconventional casing. Thus, the expansion difference (the differencebetween the respective axial thermal expansions of the casing and therotor with respect to a reference position corresponding to a thrustbearing of the turbine) is greater than that in the conventionalturbine. This results in reduction of axial clearances between therotor, i.e., a rotating body, and the components of the casing, i.e.,stationary members. Due to this, the rotor and the components of thecasing contact with each other, resulting in so-called axial-rubbing,causing the intense vibration of the shaft that hinders the continuationof the operation of the turbine.

Recently, a conventional combined-cycle steam turbine employs adouble-wall casing structure including an inner casing 111 and an outercasing 112 entirely covering turbine stages from the high-pressure firststage 7 to the high-pressure exhaust stage 8 of the high-pressuresection 5, as shown in FIG. 9, with a view to solving the foregoingproblems. This known double-wall casing structure will be referred to as“complete double-wall casing structure”, for simplicity.

Basically, thermal stress induced in a casing is proportional to thetemperature difference between the outer and inner surfaces of thecasing. Supposing that a casing is a thin-wall cylindrical structure forsimplicity, steady circumferential thermal stress due to the temperaturedifference between the outer and inner surfaces in the thin-wallcylindrical structure is expressed by: σθt=0.714α×E×T, where σθt issteady thermal stress, α is the linear thermal expansion coefficient ofthe material of the thin-wall cylindrical structure, and T is thetemperature difference between the outer and inner surfaces in thethin-wall cylindrical structure.

The temperature difference T1 between the outer and inner surfaces ofthe casing in the single-wall casing structure can be divided into0.7×T1 in the outer casing of the double-wall casing structure, and0.3×T1 in the inner casing of the same. Therefore, a steady thermalstress that will be induced in the inner casing of the double-wallcasing structure is on the order of 0.7 times a thermal stress that willbe induced in a single-wall casing structure. A steady thermal stressthat will be induced in the outer casing of the double-wall casingstructure is on the order of 0.3 times the thermal stress that will beinduced in the single-wall casing structure. Thus, the steady thermalstress induced in the casing of the high-pressure section can beeffectively reduced by using a double-wall casing structure.

Supposing that a casing is a thin-wall cylindrical structure forsimplicity, circumferential stress induced in the thin-wall cylindricalstructure due to the internal pressure therein is expressed by:σθp=a×p/t, where σθp is circumferential stress, and t is the thicknessof the thin-wall cylindrical structure. Thus, the pressure difference P1between the internal and the external pressure of the casing of thesingle-wall casing structure can be divided into 0.7×P1 for the outercasing of a double-wall casing structure, and 0.3×T1 for the innercasing of the double-wall casing structure.

Supposing that a casing is a thin-wall cylindrical structure, the radiusof an inner casing of a double-wall casing structure is about 0.9×a andthat of an outer casing of the double-wall casing structure is about1.5×a, where a is the radius of a single-wall casing. Therefore, thewall thickness of the single-wall casing is a×P1/σ1, the wall thicknessof the outer casing of the double-wall casing structure is about0.45×a×P1/σ2 and the wall thickness of the inner casing of thedouble-wall casing structure is about 0.63×a×P1/σ3, where σ1 is acircumferential pressure stresses induced in the single-wall casing, andσ2 and σ3 are circumferential pressure stresses induced in the inner andouter casings of the double-wall casing structure, respectively.

If those circumferential stresses may be equal, i.e., σ1=σ2=σ3, therespective wall thicknesses of the inner and outer casings of thedouble-wall casing structure may be about 0.63 times and about 0.45times the wall thickness of the single-wall casing, respectively.

Conversely, the respective wall thicknesses of the inner and outercasings of the double-wall casing structure may be about 0.9 times andabout 0.65 times the wall thickness of the single-wall casing,respectively, if it is desired to limit the pressure stress induced inthe double-wall casing structure to a value 0.7 times the pressurestress induced in the single-wall casing. That is, the double-wallcasing structure achieves reduction in the pressure stress whilereducing the wall thickness.

Thus, the double-wall casing structure, as compared with the single-wallcasing, is capable of reducing both steady thermal stress and pressurestress.

On the other hand, in a state such as the turbine is starting up, thetemperature of the casing rises sharply, a high thermal stress isinduced unsteadily in the casing and the casing is deformed at the sametime. The respective magnitudes of the thermal stress and the thermaldeformation are basically proportional to the temperature differencebetween the inner and outer surfaces of the casing. This temperaturedifference is greatly dependent on the wall thickness of the casing in astate where steam temperature and heat transfer coefficient changesharply, such as a state where the turbine is starting up.

The temperature difference between the inner and outer surfaces of thesingle-wall casing is large, because the inner surface of thesingle-wall casing is exposed directly to main steam and the outersurface of the same is exposed indirectly through lagging materials tothe atmosphere. On the contrary, with the double-wall casing structure,the temperature differences between the inner and outer surfaces of theouter and inner casings are smaller by far than that in the single-wallcasing. This is because, temperature difference between the inner andouter surface of the casing structure is distributed between the innerand outer casings, and temperature of steam applied to the inner andouter surfaces of the casing structure is distributed between the innerand outer casings.

Generally, the respective magnitudes of the thermal stress inducedunsteadily in the casing and the thermal deformation of the casing areproportional to the temperature difference between the inner and outersurfaces of the casing. Therefore, the thermal stress induced unsteadilyin the casings of the double-wall casing structure and the thermaldeformation of the casings of the double-wall casing structure aresmaller than those of the single-wall casing.

Steels for making the casing of a steam turbine have low thermalconductivities. Thus, if the casing has a thick wall, the conduction ofheat from the inner surface to the outer surface of the casing takes along time and the temperature difference between the inner and outersurfaces of the casing is large. In this respect, a double-wall casingstructure, in which the respective wall thicknesses of the inner andouter casings may be smaller than that of a single-wall casing, iseffective in suppressing an excessive increase of unsteady thermalstress and unsteady thermal deformation.

Since a double-wall casing structure, as compared with a single-wallcasing, reduces the temperature difference between the internal andexternal atmospheres of the casing and the wall thickness, thetemperature difference between the inner and outer surfaces of thecasing can be greatly reduced. Consequently, an excessive increase ofthermal stress and thermal deformation at the start of the turbine canbe suppressed.

As mentioned above, a double-wall casing structure, as compared with asingle-wall casing, is capable of reducing pressure stress, steadythermal stress, unsteady thermal stress and unsteady thermaldeformation. Hence, double-wall casing structure is effective inpreventing creep damage, thermal fatigue damage and damage resultingfrom high temperature low-cycle fatigue to the casing, and troubles,such as steam leakage through the horizontal flange of the casing.

However, the complete double-wall casing structure is inevitably costly.This is because, the outer casing of the complete double-wall casingstructure included in the high-pressure section of a conventionallarge-capacity, industrial steam turbine and entirely covering a part ofsteam turbine from the high-pressure first stage 7 to the high-pressureexhaust stage 8 is very large. Since the complete double-wall casingstructure has complicated construction and needs a large number of boltsfor fastening a casing-horizontal-flange joining together an upper andlower halves of the casing, assembling and disassembling the turbine forperiodic inspection or maintenance requires complicated work and a longtime. Consequently, periodic inspection and the like need increasecosts, periodic inspection needs a long time, whereby the availabilityof the power generating system decreases and power generating costincreases.

It is a still more important problem that the employment of the completedouble-wall casing structure enhances the risk of axial-rubbing. Thethermal expansion of the outer casing of the complete double-wall casingstructure is small because the temperature of steam on the inner surfaceof the outer casing is approximately equal to the temperature ofhigh-pressure exhaust steam, which is the lowest of those of steam inthe high-pressure section.

Therefore, the axial elongation difference between a rotor shaft 10which is a rotating member and a part of the casing which is astationary member, in the vicinity of a shaft seal 9 on thehigh-pressure exhaust side, is very large, as compared with such anaxial elongation difference in a single-wall casing structure. Thisresults in reduction in axial clearance. Consequently, the completedouble-wall casing structure comes into axial contact with the rotorshaft 10 to cause axial vibrations generally called rubbing vibrations.Excessively intense axial vibrations hinder the operation of the turbineand increase greatly the risk of significantly getting the reliabilityof the turbine worse.

If the axial clearance is increased to reduce such risk, the amount ofsteam leakage through the shaft seal increases to make the performanceof the turbine worse, which is undesirable from the viewpoint ofperformance. Actually, a considerably large axial clearance, as comparedwith an axial clearance required by the single-wall casing, must besecured in the shaft seal of the complete double-wall casing structure.Consequently, the leakage of steam through the shaft seal increases andmake the performance of the turbine worse.

Those problems are true of other industrial steam turbines that arerequired to operate on high-pressure, high-temperature steam as well asof steam turbines for combined-cycle power generation.

Furthermore, since the steam turbine for combined-cycle power generationis a small-capacity or a medium-capacity, the flow of main steam is lowand blade height is liable to become short and the performance of thesteam turbine is worse. Therefore, by evaluating the relation betweenthe root circle diameter and tip circle diameter of the moving bladesdesirable in respect of structural strength and performance of a steamturbine for combined-cycle power generation to enable the steam turbineto exercise satisfactory performance, the decline of the performancemust be prevented.

SUMMARY OF THE INVENTION

The present invention has been made in view of the foregoingcircumstances and it is therefore an object of the present invention tosolve problems of securing strength at elevated temperatures and ofpreventing steam leakage that arise when high-pressure, high-temperaturesteam is used for driving a steam turbine, and problems of preventingthe occurrence of rubbing due to excessive elongation difference and ofminimizing steam leakage from shaft seals.

With the foregoing object in view, the present invention provides anaxial-flow steam turbine, which includes a high-pressure sectionprovided with a turbine casing, the turbine casing having: a double-wallcasing structure, having an inner casing and an outer casing, arrangedat an area corresponding to stages from a high-pressure first stage to apredetermined high-pressure stage arranged on an upstream side of ahigh-pressure final stage; and a single-wall casing structure arrangedat an area corresponding to stages from a stage located next to saidpredetermined high-pressure stage to said high-pressure final stage.

The partial double-wall casing structure is preferably applied to asteam turbine that employs main steam having a pressure not lower than120 kgf/cm² and a temperature not lower than 550° C., and has a ratedoutput power of 120 MW or above.

It is also preferable that the double-wall casing structure is arrangedso that steam pressure in a steam passage corresponding to thedouble-wall casing structure is 90 kgf/cm² or above, or that steamtemperature in a steam passage corresponding to the double-wall casingstructure is 480° C. or above.

The present invention also provides an axial-flow steam turbine, whichincludes a high-pressure section and an intermediate-pressure section,wherein steam discharged from the high-pressure section is reheated by asteam reheater, and the steam thus reheated is supplied to theintermediate-pressure section, wherein said high-pressure section has aturbine casing having: a double-wall casing structure, having an innercasing and an outer casing, arranged at an area corresponding to stagesfrom a high-pressure first stage to a predetermined high-pressure stagearranged on an upstream side of a high-pressure final stage; and asingle-wall casing structure arranged at an area corresponding to stagesfrom a stage located next to said predetermined high-pressure stage tosaid high-pressure final stage, wherein said intermediate-pressuresection has a turbine casing having: a double-wall casing structure,having an inner casing and an outer casing, arranged at an areacorresponding to stages from an intermediate-pressure first stage to apredetermined intermediate-pressure stage arranged on an upstream sideof an intermediate-pressure final stage; and a single-wall casingstructure arranged at an area corresponding to stages from anintermediate-pressure stage located next to said predeterminedintermediate-pressure stage to said intermediate-pressure final stage,and wherein said inner casings of the said high-pressure section andintermediate-pressure section are integrated.

The partial double-wall casing structures of the high and intermediatepressure sections are preferably applied to a steam turbine that employsmain steam having a pressure not lower than 120 kgf/cm² and atemperature not lower than 550° C., and has a rated output power of 120MW or above, and wherein a temperature of reheat steam is 550° C. orabove.

It is also preferable that the double-wall casing structures of the highand intermediate pressure sections are arranged so that steamtemperature in a steam passage corresponding to the double-wall casingstructure is 480° C. or above.

In the event that the aforementioned partial double-wall casingstructure is applied, it is preferable that the outer casing is made ofa low alloy steel containing 1 to 3% Cr, such as a CrMoV alloy steel,and the inner casing is made of a Cr steel containing 8 to 10% Cr or aCr steel containing 9.5 to 12.5% Cr. Alternatively, both the outer andinner casings may be made of a low alloy steel containing 1 to 3% Cr,such as CrMoV steel.

It is preferable that, in the stages of the high-pressure sectioncorresponding to the double-wall casing structure, 0.85<Dr/Dt<0.95,where Dr is root circle diameter including roots of moving blades and Dtis tip circle diameter including tips of the moving blades.

The steam turbine provided with the partial double-wall casingstructures is suitable for use in combined-cycle power generatingsystems, thermal power plants without being combined with a gas turbine,or industrial power generating systems.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a longitudinal sectional view of principal parts of ahigh-pressure section and an intermediate-pressure section of a steamturbine in a first embodiment according to the present invention;

FIG. 2 is a graph showing the temperature dependence of the proof stressand the 10⁵-hours rapture strength of a material of a casing included ina steam turbine;

FIG. 3 is a schematic view of assistance in explaining the arrangementof a steam turbine and a generator in a conventional thermal powerplant;

FIGS. 4A and 4B are schematic views of assistance in explaining thearrangement of a gas turbine, a steam turbine and a generator in asingle-shaft combined-cycle power plant;

FIG. 5 is a side elevation of assistance in explaining the root and tipof a moving blade;

FIG. 6 is a graph showing the dependence of stress induced in a rotatingpart of a steam turbine on Dr/Dt;

FIG. 7 is a longitudinal sectional view of principal parts of ahigh-pressure section and an intermediate-pressure section of a steamturbine in a second embodiment according to the present invention;

FIG. 8 is a longitudinal sectional view of principal parts of ahigh-pressure section and an intermediate-pressure section of aconventional steam turbine employing a single-wall casing structure; and

FIG. 9 is a longitudinal sectional view of principal parts of ahigh-pressure section and an intermediate-pressure section of aconventional steam turbine employing a complete double-wall casingstructure in a high-pressure section.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

Preferred embodiments of the present invention will be described withreference to the accompanying drawings.

[First Embodiment]

A steam turbine in a first embodiment according to the present inventionwill be described with reference to FIG. 1. FIG. 1 is a longitudinalsectional view of principal parts of a high-pressure section 5 and anintermediate-pressure section 6 of the steam turbine in the firstembodiment, in which a low-pressure section is omitted.

Each of the high-pressure section 5 and the intermediate-pressuresection 6 has a plurality of stages each consisting of a combination ofstationary blades 3 and moving blades 4. The moving blades 4 of thehigh-pressure section 5 and the intermediate-pressure section 6 areattached to a common rotor shaft 10.

Main steam flows through an inlet port 5 a into the high-pressuresection 5, acts on a high-pressure first stage 7, flows sequentiallythrough the rest of the high-pressure stages, leaves a high-pressureexhaust stage 8 and exhausts through an outlet port 5 b. The steamexhausted through the outlet port 5 b flows through an inlet port 6 ainto the intermediate-pressure section 6, acts on anintermediate-pressure first stage 12, flows sequentially through therest of the intermediate-pressure stages, leaves anintermediate-pressure exhaust stage 13 and exhausts through an outletport 6 b.

Shown also in FIG. 1 are shaft seals 9 and casing attachments 11.

As shown in FIG. 1, the high-pressure section 5 has a double-wall casingstructure, which is composed of an inner casing 1 and an outer casing 2.The double-wall casing structure is arranged at an area corresponding tothe stages from the high-pressure first stage 7 to a predeterminedhigh-pressure stage on the upstream side of the high-pressure exhauststage 8 (from the high-pressure first stage 7 to a high-pressure fourthstage, with the embodiment of FIG. 1).

A single-wall casing structure, which has only the outer casing 2, isarranged at an area corresponding to the stages which follow saidpredetermined stage (high-pressure fourth stage with the embodiment ofFIG. 1), in other words, from the high-pressure fifth stage to thehigh-pressure exhaust stage 8. As mentioned above, the casing of thehigh-pressure stage 5 has a “partial double-wall casing structure.” Asshown in FIG. 1, the outer casing 2 is formed continuously from the partcorresponding to the high-pressure first stage 7 to the partcorresponding to the high-pressure exhaust stage 8.

The high-pressure section 5 provided with the partial double-wall casingstructure is preferably applied to steam turbines driven by main steamhaving a pressure of 120 kgf/cm² or above and a temperature of 550° C.or above, and having a rated output power of 120 MW or above.Preferably, the high-pressure section 5 is provided with the double-wallcasing structure at an area where steam pressure in a steam passage is90 kgf/cm² or above, or steam temperature in the steam passage is 480°C. or above.

The reason for determining the area provided with the double-wall casingstructure is as follows. Generally, the creep rate of materials of acasing of a steam turbine increases remarkable at temperatures exceeding480° C. Therefore, the reduction of high-temperature strength due tocreep must be taken into consideration in designing the casing. FIG. 2is a graph showing the temperature dependence of the proof stress andthe 10⁵-hour rapture strength of a material of a casing included in asteam turbine, in which stress S is measured on the vertical axis andtemperature T is measured on the horizontal axis. As shown in FIG. 2,the proof stress varies with temperature along a broken line B-B′, andthe 10⁵-hour rapture strength varies along a continuous line A-A′. Thebroken line B-B′ and the continuous line A-A′ intersect each other at apoint P substantially corresponding to 480° C.

The proof stress is used as a design criterion for a temperature rangebelow about 480° C., and the 10⁵-hour rapture strength is used as adesign criterion for a temperature range above about 480° C. Thus, acurve B-P-A′ is used for finding a reference material strength indesigning the casing. Therefore, the double-wall casing structure isemployed at an area corresponding to stages exposed to heat oftemperatures in the temperature range in which the strength of materialsdecreases sharply, in other words, stages exposed to heat oftemperatures in a temperature range above 480° C. in which creep rupturestrength must be used as a design criterion, to cope with the sharpreduction of material strength at high temperatures effectively.

The Plandtl number of steam has a significant influence on heat transfercoefficient. In a steam turbine used in a conventional thermal powerplants and combined-cycle power plants for general thermal powergeneration, the Plandtl number of steam in a steam passage is about 1.0,and the steam has a temperature of about 480° C. and a pressure on theorder of 90 kgf/cm².

Therefore, if an area of the casing corresponding to a steam passage inwhich the pressure of the steam is 90 kgf/cm² or above, or correspondingto a steam passage in which the temperature of steam is 480° C. orabove, is provided with the double-wall casing structure, thermal stressinduced in the casing and the axial elongation difference can be limitedwithin design allowance ranges with sufficient tolerances, and thethermal deformation of the casing can be limited to a satisfactorily lowlevel. Thus, a highly reliable, safe steam turbine free of damage andsteam leakage that will hinder the continuation of operation can beprovided.

As mentioned above, the high-pressure section of the steam turbine isprovided with the double-wall casing structure at an area of the casingto be exposed to high-pressure, high-temperature steam, therebysuppressing the induction of excessively high thermal stress and thedevelopment of excessively large thermal deformation. In addition, thedouble-wall casing structure is not provided up to the high-pressureexhaust stage, in other words, the area where the double-wall casingstructure is arranged is limited, thereby preventing the excessiveincrease of the axial elongation difference. Consequently, problems ofensuring high strength at high temperatures and reducing steam leakageinvolved in raising the temperature and pressure of steam for drivingthe steam turbine can be solved. Rubbing can be prevented by suppressingthe development of excessive elongation difference and a safe steamturbine free of vibrations that hinder the operation of the steamturbine can be provided, and increase in manufacturing cost andoperating cost can be suppressed.

Materials of the inner casing 1 and the outer casing 2 will be describedhereinafter.

In the high-pressure section 5 of the steam turbine provided with thepartial double-wall casing structure as shown in FIG. 1, it ispreferable that the outer casing 2 is made of a low alloy steelcontaining 1 to 3% Cr, such as a CrMoV steel, and the inner casing 1 ismade of a 9-Cr steel containing 8 to 10% Cr or a 12-Cr steel containing9.5 to 12.5% Cr.

Reduction in manufacturing cost is achieved, by using the 12-Cr steel orthe 9-Cr steel having a high high-temperature strength only for theinner casing 1 to be exposed to high-pressure, high-temperature steam.Increase in the axial elongation difference can be suppressed by formingonly the specific part of the casing of the 12-Cr steel or the 9-Crsteel having a small thermal expansion coefficient. Consequently,increase in steam leakage through the shaft seals 9 due to big axialclearances can be prevented, and the risk of generating axial vibrationsby axial rubbing can be reduced. Thus, the steam turbine can bemanufactured at a low cost and can operate at a low operating cost.

The partial double-wall casing structure is far less subject to thermaldeformation, and a thermal stress induced in the partial double-wallcasing structure is far less than that induced in a single-wall casingstructure. Therefore, both the inner casing 1 and the outer casing 2 maybe made of a low alloy steel containing 1 to 3% Cr, typically CrMoVsteel. Although the partial double-wall casing structure of such a lowalloy steel requires careful consideration in designing the same,increase in the cost can be limited to the least extent, and the axialelongation difference is small. Therefore, steam leakage through theshaft seals can be minimized and axial rubbing can be effectivelyprevented.

Preferably, the moving blades 4 of the stages, which correspond to thearea where the double-wall casing structure is employed, of thehigh-pressure section 5 meet an inequality: 0.85<Dr/Dt<0.95, where Dr isroot circle diameter of the moving blades 4, and Dt is tip circlediameter of the moving blades 4. The reasons for this condition will bedescribed hereinafter with reference to FIGS. 3 to 6.

Generally, the diameter of a rotor shaft 14 in a high-pressure sectionof a steam turbine for a combined-cycle plant is greater than that ofthe rotor shaft of a conventional steam turbine for thermal powergeneration equivalent in capacity thereto for the following reasons.

Referring to FIG. 3, a general, conventional thermal power plant has asteam turbine 15 and a generator 16. The diameter of a rotor shaft 14included in a high-pressure section of the steam turbine 15 does notneed to be big because the rotor shaft 14 needs to transmit only shafttorque generated by the high-pressure section of the steam turbine 15.

Recently, in a combined-cycle plant, a single-shaft power generatingsystem formed by coaxially arranging g a gas turbine 17, a steam turbine15 and a generator 16 as shown in FIG. 4 has been generally used. In thearrangement shown in FIG. 4, a shaft torque of the rotor shaft 15 of thehigh-pressure section of the steam turbine 15, and a shaft torque of thegas turbine 17 are used in combination. Therefore, the rotor shaft 14 ofthe high-pressure section of the steam turbine 15 must have a bigdiameter to have a necessary torsional strength.

Referring to FIG. 5, if the diameter of the rotor shaft 14 of the steamturbine 15 is increased, the diameter of a circle including the roots ofmoving blades 4 increases accordingly. However, since the flow rate doesnot change, the blade height 43 of the moving blades 44 must beinevitably reduced to maintain the exit area of the moving bladessubstantially constant.

Supposing that the moving blade height 43 is Hb and the moving bladewidth 44 is Wb, the influence of a secondary flow in flows through thecascade increases sharply and the hydrodynamic performance of the movingblades deteriorates sharply when Hb/Wb<1. Therefore, falling into such acondition must be avoided.

Thus, it is desirable to increase the blade height of the moving bladesof the high-pressure section having a short blade height, and it isdesirable that Dr/Dt is small.

Generally, the blade width Wb of moving blades of the high-pressuresection of a steam turbine for a single-shaft combined-cycle power planthaving an output power of 120 MW or above is on the order of 20 mm atthe minimum, and the root circle diameter Dr is about 800 mm at thesmallest because it is difficult to reduce the root circle diameter Drgreatly. Therefore, it is important, for maintaining the performance ofthe steam turbine on a high level, to satisfy conditions expressed by:$\begin{matrix}{{{Dr}/{Dt}} < {\left( {{Dt} - {2{Hb}}} \right)/{Dt}}} \\{{= {1 - {2{{Hb}/{Dt}}}}}} \\{{\approx {1 - {2{{Wb}/\left( {{Dr} + {2{Wb}}} \right)}}}}} \\{{\approx {1 - {2 \times {20/\left( {800 + {2 \times 20}} \right)}}}}} \\{{\approx 0.95}}\end{matrix}$

accordingly, Dr/Dt<0.95.

The high-pressure section of the steam turbine is exposed to hightemperatures. In most cases, the stages of the high-pressure sectioncorresponding to the double-wall casing structure are exposed to hightemperatures not lower than 480° C. Therefore, stresses are induced inthe materials of the moving blades and the rotor in a mode as shown inFIG. 2, and the reduction of the strength of the moving blades and therotor at high temperatures is a significant problem. If the movingblades have an excessively big height, the moving blades are undergocreep damage and the probability of breakage of the moving blades or therotor wheel while the steam turbine is in operation increases sharply.

Generally, a steam turbine is designed such that a local stress inducedin a moving blade holding part 411 and a stress induced in a centralpart of a rotor shaft 14 are on the substantially same level. Since theroot circle diameter Dr of the moving blades of a steam turbine isdetermined so that the performance of the steam turbine andmanufacturing techniques are matched, the root circle diameters Dr ofcomparatively large steam turbines having an output power of 120 MW orabove are not greatly different from each other.

Supposing that Dr is fixed, a local stress induced in a moving bladeholding part and a circumferential stress induced in a central part of arotor shaft decreases as Dr/Dt increases as shown in FIG. 6. Whereas thelocal stress induced in the moving blade holding part decreases sharply,the circumferential stress induced in the central part of the rotorshaft decreases gradually and changes scarcely in a range where Dr/Dt ishigh.

Generally, the circumferential stress induced in the central part of therotor shaft remains at a level near the limit of strength of the rotorshaft in the high Dr/Dt range. Therefore, the design and manufacture ofthe rotor shaft on the basis of a stress far higher than such acircumferential stress are impossible. In a low Dr/Dt range, the localstress induced in the moving blade holding part exceeds thecircumferential stress induced in the central part of the rotor shaftand increases sharply as the Dr/Dt decreases. Therefore the design andmanufacture of the rotor shaft on the basis of data in such a range aredifficult.

Referring to FIG. 6, it is known empirically that a curve 62 indicatingthe variation of the local stress with Dr/Dt and a curve 61 indicatingthe variation of circumferential stress with Dr/Dt intersect each otherat a point corresponding a Dr/Dt of about 0.85. In a Dr/Dt range below0.85, stress exceeds limit strength and hence a rotor having a Dr/Dtbelow 0.85 is unrealizable. Therefore, a steam turbine must meet:0.85<Dr/Dt in view of the high-temperature strength of the rotating partof the steam turbine.

As obvious from the foregoing description, when the stages correspondingto the double-wall casing structure of the high-pressure section meet:0.85<Dr/Dt<0.95, where Dr is root circle diameter and Dt is tip circlediameter, the deterioration of the performance of moving blades by theeffect of the secondary flow can be prevented, the performance of thehigh-pressure section of the steam turbine can be maintained at a highlevel, and a highly reliable, safe steam turbine free of damage in themoving blades or the rotor wheel that may result in breakage can beprovided.

[Second Embodiment]

A steam turbine in a second embodiment according to the presentinvention will be described with reference to FIG. 7. FIG. 7 is alongitudinal sectional view of principal parts of a high-pressuresection 5 and an intermediate-pressure section 6 of the steam turbine inthe second embodiment. In FIG. 7, a low-pressure section is omitted. InFIG. 7, parts like or corresponding to those of the first embodiment aredenoted by the same reference characters and the description thereofwill be omitted to avoid duplication.

The steam turbine in the second embodiment is a reheat cycle steamturbine that reheats steam discharged through an outlet port 5 b of ahigh-pressure section 5 by a steam reheater, not shown, and supplies thereheat steam to an intermediate-pressure section 6 through an inlet port6 a.

The steam turbine in the second embodiment is suitable for use as asteam turbine using main steam of a pressure not lower than 120 kgf/cm²and a temperature not lower than 550° C., and having a rated outputpower of 120 MW or above.

As shown in FIG. 7, with the steam turbine in the second embodiment,similarly to that in the first embodiment, the high-pressure section 5has a double-wall casing structure, which is composed of an inner casing101 and an outer casing 102. The double-wall casing structure isarranged at an area corresponding to the stages from the high-pressurefirst stage 7 to a predetermined high-pressure stage on the upstreamside of the high-pressure exhaust stage 8 (from the high-pressure firststage to a high-pressure fourth stage, with the embodiment of FIG. 7). Asingle-wall casing structure, which has only the outer casing 102, isarranged at an area corresponding to the stages which follow saidpredetermined stage.

With the steam turbine in the second embodiment, theintermediate-pressure section 6 also has a double-wall casing structureat an area corresponding to the stages from the intermediate-pressurefirst stage 12 to a predetermined intermediate-pressure stage on theupstream side of the intermediate-pressure exhaust stage 13 (from theintermediate-pressure first stage to a intermediate-pressure secondstage, with the embodiment of FIG. 7). The steam turbine has asingle-wall casing structure at an area corresponding to the stageswhich follow said predetermined stage, in other words, from theintermediate-pressure third stage to the intermediate-pressure exhauststage. Thus, both the high-pressure section 5 and theintermediate-pressure section 6 are provided with the partialdouble-wall casing structures.

As shown in FIG. 7, the inner casing 101 is formed continuously from thepart corresponding to the high-pressure fourth stage to the partcorresponding to the intermediate-pressure second stage. Thus, the innercasing 101 covers both the high-pressure section 5 and theintermediate-pressure section 6, and is integrally formed for both thehigh-pressure section 5 and the intermediate-pressure section 6.Similarly, the outer casing 102 covers both the high-pressure section 5and the intermediate-pressure section 6, and is integrally formed forboth the high-pressure section 5 and the intermediate-pressure section6.

Since high-temperature, high-pressure steam is supplied also to theintermediate-pressure section 6 of the steam turbine in the secondembodiment, i.e., the reheat cycle steam turbine, theintermediate-pressure section 6 is provided with the partial double-wallcasing structure. Thus, the second embodiment is substantially the samein operation and effect as the first embodiment.

The area of the intermediate-pressure section 6 provided with thedouble-wall casing structure may be determined on the basis of ideasexplained previously in connection with the first embodiment. It istherefore preferable to arrange the double-wall casing structure at anarea where the pressure of steam in the steam passage is 90 kgf/cm² orabove with the double-wall casing structure, or where the temperature ofsteam in the steam passage is 480° C. or above.

The materials of the inner casing 101 and the outer casing 102 may beselectively determined on the basis of ideas previously explained inconnection with the first embodiment. The outer casing 102 may be madeof a low alloy steel containing 1 to 3% Cr represented by a CrMoV alloysteel, and the inner casing 101 may be made of a 9-Cr steel containing 8to 10% Cr or a 12-Cr steel containing 9.5 to 12.5% Cr. Both the innercasing 101 and the outer casing 102 may be made of a low alloy steelcontaining 1 to 3% Cr represented by a CrMoV alloy steel.

The steam turbine provided with the foregoing partial double-wall casingstructure (both the first and the second embodiments) is suitable foruse in a combined-cycle power generating system including a gas turbineand a steam turbine. The steam turbine of the present invention isapplicable to a combined-cycle power generating system including asteam-cooled gas turbine cooled by using steam. The steam turbineprovided with the foregoing partial double-wall casing structure can beused in a thermal power plant that does not use the steam turbine incombination with a gas turbine or can be used in an industrial thermalpower plant.

The use of the foregoing steam turbine in a thermal power plantsuppresses the increase of the operating cost of the power plantoperating under high steam conditions including high pressure and hightemperature. The foregoing steam turbine exercises the same effect andoperation not only when the same is used in a combined-cycle powerplant, but also when the same is used in a thermal power plant not usingthe steam turbine in combination with a gas turbine or an industrialpower generating system, and when raising the pressure and temperatureof steam to be used.

What is claimed is:
 1. An axial-flow steam turbine comprising ahigh-pressure section provided with a turbine casing, said turbinecasing having: a double-wall casing structure, having an inner casingand an outer casing, arranged at an area corresponding to a part from aninlet part of said high-pressure stage to a predetermined high-pressurestage arranged on an upstream side of a high-pressure final stage; and asingle-wall casing structure arranged at an area corresponding to stagesfrom a stage located next to said predetermined high-pressure stage tosaid high-pressure final stage.
 2. The steam turbine according to claim1, wherein said steam turbine employs main steam having a pressure of120 kgf/cm² or above and a temperature of 550° C. or above, and has arated output power of 120MW or above.
 3. The steam turbine according toclaim 1 or 2, wherein said double-wall casing structure is arranged sothat steam pressure in a steam passage corresponding to the double-wallcasing structure is 90 kgf/cm² or above, or that steam temperature in asteam passage corresponding to the double-wall casing structure is 480°C. or above.
 4. An axial-flow steam turbine comprising a high-pressuresection and an intermediate-pressure section, wherein steam dischargedfrom said high-pressure section is reheated by a steam reheater, and thesteam thus reheated is supplied to said intermediate-pressure section,wherein said high-pressure section has a turbine casing having: adouble-wall casing structure, having an inner casing and an outercasing, arranged at an area corresponding to a part from an inlet partof said high-pressure stage to a predetermined high-pressure stagearranged on an upstream side of a high-pressure final stage; and asingle-wall casing structure arranged at an area corresponding to stagesfrom a stage located next to said predetermined high-pressure stage tosaid high-pressure final stage, wherein said intermediate-pressuresection has a turbine casing having: a double-wall casing structure,having an inner casing and an outer casing, arranged at an areacorresponding to a part from an inlet part of said intermediate-pressurestage to a predetermined intermediate-pressure stage arranged on anupstream side of an intermediate-pressure final stage; and a single-wallcasing structure arranged at an area corresponding to stages from anintermediate-pressure stage located next to said predeterminedintermediate-pressure stage to said intermediate-pressure final stage,and wherein said inner casings of the said high-pressure section andintermediate-pressure section are integrated.
 5. The steam turbineaccording to claim 4, wherein said steam turbine employs main steamhaving a pressure of 120 kgf/cm² or above and a temperature of 550° C.or above, and has a rated output power of 120 MW or above, and wherein atemperature of the reheat steam is 550° C. or above.
 6. The steamturbine according to claim 4 or 5, wherein in said high-pressure sectionand said intermediate-pressure section, said double-wall casingstructure is arranged so that steam temperature in a steam passagecorresponding to the double-wall casing structure is 480° C. or above.7. The steam turbine to any one of claims 1 to 6, wherein said outercasing are made of a low alloy steel containing 1 to 3% Cr, such as aCrMoV alloy steel, and said inner casing are made of a Cr steelcontaining 8 to 10% Cr or a Cr steel containing 9.5 to 12.5% Cr.
 8. Thesteam turbine according to any one of claims 1 to 6, wherein both theouter and inner casings are made of a low alloy steel containing 1 to 3%Cr, such as CrMoV steel.
 9. The axial-flow steam turbine according toany one of claims 1 to 8, wherein, in said stages of the high-pressuresection corresponding to said double-wall casing structure,0.85<Dr/Dt<0.95, where Dr is root circle diameter including roots ofmoving blades and Dt is tip circle diameter including tips of the movingblades.
 10. A combined-cycle power generating system comprising a gasturbine and the axial-flow steam turbine according to any one of claims1 to
 9. 11. The combined-cycle power generating system according toclaim 10, wherein the gas turbine is of a steam-cooled type cooled byusing steam.